Systems and Methods for Implementing Ejector Refrigeration Cycles with Cascaded Evaporation Stages

ABSTRACT

Systems and methods for implementing ejector refrigeration cycles with cascaded evaporation stages that utilize a pump to optimize operation of the ejector and eliminate the need for a compressor between the evaporation stages.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is continuation of U.S. application Ser. No.17/766,796, which is a U.S. National Stage Application of PCT PatentApplication Serial No. PCT/US20/62972, which claims priority to U.S.Provisional Application No. 62/943,542, each of which are incorporatedherein by reference. This application and U.S. Pat. Nos. 10,514,201,10,533,793, 10,465,983 and 10,514,202, which are incorporated herein byreference, are commonly assigned to Bechtel Energy Technologies &Solutions, Inc.

FIELD OF THE DISCLOSURE

The present disclosure generally relates to systems and methods forimplementing ejector refrigeration cycles with cascaded evaporationstages. More particularly, the present disclosure utilizes a pump tooptimize operation of the ejector and eliminate the need for acompressor between the evaporation stages.

BACKGROUND

Ejector refrigeration cycles offer performance advantages compared totraditional cascaded refrigeration cycles. One example of a conventionalejector refrigeration cycle system 100 with two evaporation pressures isillustrated by the schematic diagram in FIG. 1 .

In FIG. 1 , condensed liquid refrigerant in line 102 is transferred tothe motive nozzle inlet of an ejector 104 where it is expanded andaccelerated to a lower pressure and partially vaporized. Vaporizedrefrigerant from a first evaporative heat exchanger 116 is transferredthrough line 118 to the suction nozzle of the ejector 104. Through theentrainment of suction flow in line 118 by the acceleration of themotive flow in line 102, the ejector 104 discharges these two flows at apressure higher than the pressure in the heat exchanger 116 in atwo-phase state after mixing and diffusing, resulting in a smallerpressure ratio across a compressor 112 and thus, reducing the powerconsumption of the compressor 112. The partially vaporized flow iscollected in a flash tank 108 utilized for phase separation of thetwo-phase flow. Vapor in line 110 from the flash tank 108 isrecompressed in the compressor 112 and subsequently condensed in acondensing heat exchanger 114 with heat rejection to an external heatsink. Liquid from the flash tank 108 is evaporated in the heat exchanger116 used to cool an external stream. An additional chilling stage isachieved by expanding liquid from the flash tank 108 to a low-pressurestage in line 120 where the flow is evaporated in a second evaporativeheat exchanger 124. Vaporized refrigerant is then compressed to higherstages in a booster compressor 122.

A pressure differential between the motive flow in line 102 upstream ofthe ejector 104 and the suction flow in line 118 is what allows theejector 104 to operate. A certain difference in pressure between thehigh side (motive) and low side (suction) is needed for the ejector 104to perform efficiently. Because of this, many manufacturers will usemultiple ejectors in parallel or an ejector that is adjustable toaccommodate for changes in the upstream motive pressure. As a result,high-side pressure control is a limiting aspect that makes ejectorrefrigeration cycles less suitable for operational flexibility over awide range of operating conditions that can impact the performance ofthe ejector, leading to a significantly reduced coefficient ofperformance. Additionally, to achieve two evaporation pressures, eithera booster compressor between the two evaporator heat exchangers 116 and124 would be required or the medium temperature at the outlet of heatexchanger 116 would need to be expanded down to the lower evaporatingpressure at the outlet of heat exchanger 124. The addition of a boostercompressor is more efficient than expanding the outlet flow from heatexchanger 116 because a single compressor would have a higher-pressureratio to overcome, and thus, consume more power. Because the additionalbooster compressor 122 also increases the overall power consumption ofthe cycle, alternative means of compressing the working fluid to achievethe pressure differential between the heat exchangers 116 and 124 may bepreferred and multiple ejectors may be required to maintain optimaloperational performance and energy efficiency.

BRIEF DESCRIPTION OF THE DRAWINGS

The present disclosure is described below with references to theaccompanying drawings in which like elements are referenced with likenumerals and which:

FIG. 1 is a schematic diagram illustrating a conventional ejectorrefrigeration cycle system.

FIG. 2 is a schematic diagram illustrating an embodiment of an improvedejector refrigeration cycle system with two evaporation pressures.

FIG. 3 is a schematic diagram illustrating an embodiment of an improvedejector refrigeration cycle system with two evaporation pressures andsensors for adjustable operation.

FIG. 4 is a P-h diagram of the ejector refrigeration cycle systemillustrated in FIG. 2 using carbon dioxide (R744) as the refrigerant.

FIG. 5 is a P-h diagram of the ejector refrigeration cycle systemillustrated in FIG. 2 using propane (R290) as the refrigerant.

FIG. 6 is a P-h diagram of the ejector refrigeration cycle systemillustrated in FIG. 2 using R410A as the refrigerant.

DETAILED DESCRIPTION OF THE ILLUSTRATIVE EMBODIMENTS

The subject matter of the present disclosure is described withspecificity, however, the description itself is not intended to limitthe scope of the disclosure. The subject matter thus, might also beembodied in other ways, to include different structures, steps and/orcombinations similar to and/or fewer than those described herein, inconjunction with other present or future technologies. Although the term“step” may be used herein to describe different elements of methodsemployed, the term should not be interpreted as implying any particularorder among or between various steps herein disclosed unless otherwiseexpressly limited by the description to a particular order. Otherfeatures and advantages of the disclosed embodiments will be or willbecome apparent to one of ordinary skill in the art upon examination ofthe following figures and detailed description. It is intended that allsuch additional features and advantages be included within the scope ofthe disclosed embodiments. Further, the illustrated figures are onlyexemplary and are not intended to assert or imply any limitation withregard to the environment, architecture, design, or process in whichdifferent embodiments may be implemented. All streams described arecarried by physical lines. To the extent that temperatures and pressuresare referenced in the following description, those conditions are merelyillustrative and are not meant to limit the disclosure.

The systems and methods disclosed herein thus, improve conventionalejector refrigeration cycles with the addition of a pump positioneddownstream from the condensing heat exchanger to eliminate the need fora booster compressor and achieve the pressure differential between theevaporative heat exchangers.

In one embodiment, the present disclosure includes a refrigerationsystem, which comprises: i) an ejector in fluid communication with afirst evaporative heat exchanger, a pump and a flash drum, the ejectorpositioned downstream of the first evaporative heat exchanger and thepump and positioned upstream of the flash drum; ii) a second evaporativeheat exchanger in fluid communication with the flash drum and a singlecompressor; iii) the compressor in fluid communication with the flashdrum and positioned downstream of the flash drum; and iv) a condensingheat exchanger in fluid communication with the compressor and the pump,the condensing heat exchanger positioned downstream of the compressorand upstream of the pump.

In another embodiment, the present disclosure includes a refrigerationmethod, which comprises: i) pumping a condensed liquid refrigerant froma first heat exchanger to an ejector at a higher pressure than apressure at the first heat exchanger; ii) ejecting the liquidrefrigerant from the ejector as a two-phase refrigerant to a flash drum;iii) separating the two-phase refrigerant in the flash drum into aliquid refrigerant and a vapor refrigerant; iv) transferring a portionof the liquid refrigerant from the flash drum through a second heatexchanger to the ejector as an evaporated refrigerant with anevaporation pressure; v) transferring another portion of the liquidrefrigerant from the flash drum through a third heat exchanger to acompressor as a vaporized refrigerant at another evaporation pressure;and vi) transferring a compressed refrigerant from the compressor to thefirst heat exchanger.

In FIG. 2 , a schematic diagram illustrates an embodiment of an improvedejector refrigeration cycle system 200 with two evaporation pressures. Acondensed liquid refrigerant in line 102 from the condensing heatexchanger 114 is transmitted to a refrigerant pump 202. The condensedliquid refrigerant is pumped to a higher pressure, with pressure ratioson the order of about 1.14 to about 5 and in accordance with anacceptable entrainment ratio, defined as the ratio of suction nozzlemass flow rate to the motive nozzle mass flow rate, for the ejector 104.The preferred acceptable entrainment ratio is greater than about 0.3 tomaximize ejector efficiency and system capacity. An electronic expansionvalve 204 downstream of the condensing heat exchanger 114 ensuresoperational stability by facilitating the ability to bypass the ejector104 and helps control the pressure in the condensing heat exchanger 114.Refrigerant is transmitted from the ejector 104 through line 106 to theflash tank 108. Following phase separation in the flash tank 108, theliquid flows from the bottom of flash tank 108 are expanded to twodifferent evaporation pressures corresponding with heat exchangers 116and 124. One level is obtained through a medium temperature evaporativeheat exchanger 116. The second level is obtained at the lowesttemperature evaporative heat exchanger 124. The evaporated refrigerantfrom the evaporative heat exchanger 116 is transferred to the compressor112 through line 118. The vaporized refrigerant from the low temperatureevaporative heat exchanger 124 is transmitted to the suction nozzle ofthe ejector 104 through line 126.

The ejector refrigeration cycle system in FIG. 2 is most beneficialoperating in cooling mode due to increased pressure differentials acrossthe system resulting in larger amounts of available expansion workrecovery. As a result, heat rejection would occur in two ways. The firstway would be to reject heat to the atmosphere using the condensing heatexchanger 114. The second way would be to reject heat to a working fluidas done in a hydronic heating system through the condensing heatexchanger 114. In most applications, the pressure in the condensing heatexchanger 114 of the cycle will be dictated by the temperature of theheat sink to which heat is rejected. As such, the potential efficiencybenefits of the ejector refrigeration cycle system in FIG. 2 are higherin climates where the ambient condition is extreme and/or fluctuatesbroadly over a range of conditions. The heat sources to evaporative heatexchangers 116 and 124 may come in the form of air or a secondary waterloop. This water loop will reject heat to the vapor compression cycleand circulate through the building absorbing loads from room units,which can be controlled individually through local fan speeds. Thiswater loop acts as a buffer for the vapor compression cycle from thermalexcitations in rooms and other areas due to its significant thermalinertia. This reduction in vapor compression system excitation is abenefit for efficiency, as system stability almost always results inmore efficient operation. Furthermore, the complicated fluid and gasdynamics that occur within a two-phase ejector can be susceptible toinstabilities, making this advantage even more notable.

Utilizing a pump as a method to control the ejector inlet conditionswill enable a fixed-geometry ejector to operate at or near its designcondition over a range of operating conditions. While researchers havebeen unable to develop a variable-geometry ejector with an isentropicefficiency competitive to that of a fixed-geometry ejector, off-designoperating conditions impart inefficiencies as well. As such, the ejectorrefrigeration cycle system in FIG. 2 addresses both issues. Provided thepump 202 has a higher isentropic efficiency than the compressor 112within the system and is able to modulate the ejector 104 inletcondition to an optimum, this design will provide an increase inefficiency. As the isentropic efficiency of the pump 202 and thecompressor 112 are largely a function of pressure ratio, the combinationcould be selected and designed to operate over a smaller range ofpressure ratios, despite being over a broader range of ambienttemperatures than standard applications of either component, increasingthe possibility of higher isentropic efficiencies and thus, increasedsystem efficiency.

Referring now to FIG. 3 , a schematic diagram illustrates an embodimentof an improved ejector refrigeration cycle system 300 with twoevaporation pressures and sensors for adjustable operation. Theelectronic expansion valve 204 is modulated to establish a stablepressure differential between the medium temperature evaporative heatexchanger 116 and the condensing heat exchanger 114, as well asmodulating the liquid level in the flash tank 108. A target compressorsuction superheat of 10 K at the compressor 112 suction port will bemaintained through modulation of valves 210 and 208. The saturated vaporflow from the flash tank 108 will mix with the medium temperatureevaporator outlet flow in line 118, which will decrease and possiblyeliminate superheat at the suction port to the compressor 112. Thisneeds to be actively rectified via decreasing the orifice diameter ofvalve 208 and increasing the orifice diameter of valve 210. If thesystem is designed and charged correctly, this will represent the designoperating condition for the compressor 112 and will also result insubcooled liquid at the outlet of the condensing heat exchanger 114 of 5K. These two temperature differentials (superheat and subcool) result inthe target balance between utilization of a condenser and evaporatorsurface area, as well as compressor performance, which is consistentwith standard refrigeration applications. At the outlet of thecondensing heat exchanger 114, flow will be flowing through the ejector104 motive nozzle, which will also entrain flow from the lowertemperature evaporative heat exchanger 124 through line 126. Thepressure lift across the ejector 104 will then facilitate thedifferential in pressure between the two evaporative heat exchangers116, 124. This is the goal of the system 300. To maximize this pressuredifferential, the electronic expansion valve 204 should be closedcompletely during ejector operation.

Ideally, the system 300 requires minimal temperature and pressuremeasurement feedback to minimize cost and complexity, such that thesensors focus solely on retaining compressor suction superheat. However,in order to ensure that the pump 202 and ejector 104 combination isoperating at the optimal condition, the lower temperature evaporator 124outlet state and the condensing heat exchanger 114 outlet state need tobe considered as well. There is no need for a flow rate measurement inthe system 300, as the compressor maps can be utilized in conjunctionwith sensors and power draw measurement that will already be in place topredict the mass flow rate. To control the flow rates in the system 300,both the compressor 112 and the pump 202 will have a dedicated variablefrequency drive (VFD) to modulate their operating frequency, enablingpart-load operation when the cooling capacity required of the system 300is less than the design point. The only other method of active controlwithin the system 300 will be modulation of electronic expansion valvesas previously discussed.

The electronic expansion valve 204 may be used during startup as abypass to the ejector 104 to ensure control of the system 300, and maybe closed during steady operation to allow 100% of the condensing heatexchanger 114 outlet mass flow to enter the ejector 104 to maximize theejector 104 work recovery. The electronic expansion valve 208 betweenthe flash tank 108 and the medium temperature evaporative heat exchanger116 will control the outlet superheat temperature in line 118.Controlling the outlet superheat temperature in line 118 using flowcontrol valve 208 will contribute to control of the superheat at thecompressor suction port, which is measured by pressure sensor 344 andsensor 346 tuned with the thermo-physical properties of the primaryworking fluid to represent superheat. Valve 210 may work in tandem withvalve 208 due to the mixing of flows in line 110 and line 118 beforeentering compressor suction to maintain compressor suction superheat andto ensure an equal pressure drop across valves 208 and 210 so thatsignificant variation in flash tank 108 liquid level does not occur.Valves 208 and 210 may work with valve 206 and electronic expansionvalve 204 to ensure that the liquid level in the flash tank 108 remainsat a safe level. The liquid level can be assessed by the rate of changeof superheat values at the outlets of the evaporative heat exchangers116, 124.

For example, if the superheat values increase rapidly, this suggeststhat the liquid in the flash tank 108 has disappeared, thussignificantly increasing the inlet flow quality at the evaporative heatexchangers 116, 124. In this case, the valve 210 between the flash tank108 and the compressor 112 will open and in tandem the valve 204 willopen to pull more sub-cooled liquid from the condensing heat exchanger114 into the flash tank 108. The electronic expansion valve 206 betweenthe flash tank 108 and the low temperature evaporative heat exchanger124 will control the outlet superheat and receives inputs of thetemperature and pressure from sensor 342 and 340, respectively, at theoutlet of the low temperature evaporative heat exchanger 124, just asthe compressor suction superheat is measured. The low temperatureevaporative heat exchanger 124 outlet state will have a target superheatof 5 K to balance high ejector efficiency while retaining as high of anevaporation pressure as possible for a given heat source temperature. Acombination of the sensors 332, 334, 336, 338, 340, 342, and/or 344, 346and the compressor power may be used to control the pump speed. Giventhat the fluid on the outlet of the low temperature evaporative heatexchanger 124 represents the ejector 104 suction nozzle inlet state andthe pressure measured by sensor 336 represents the motive nozzle inletpressure, the outlet quality of the ejector 104 can be calculated viaempirically derived relations. Conservation of mass is then appliedusing the compressor 112 flow rate, and all operating parameters of theejector 104 are known, allowing the pump 202 to act to vary the motiveinlet state to an optimal condition.

For safe initialization of pump 202, the power measurement of compressor112 in conjunction with superheat from sensors 344 and 346 can be usedto calculate the mass flow rate passing through the condensing heatexchanger 114. Sensors 332 and 334 can then be used to calculate thedensity of the pump 202 inlet state, which can then be coupled with themass flow rate calculation and the known pump displacement volume todetermine an initial pump speed. This pump speed would be as close tomatching the mass flow rate of the system 300 as possible to minimizecycle excitation as well as the chance of cavitation in the pump 202suction port. Upon starting the pump 202, the calculated speed should beentered into the VFD, and immediately following the pump startup thevalve 204 should be completely closed to minimize the chance of a bypassoccurring. Once operational, the pump speed is directly proportional topump 202 discharge pressure and the two-phase quality at the ejector 104diffuser outlet in line 106. The pump discharge pressure can be measuredwith sensor 336 and if a maximum pressure is exceeded the valve 204should be opened and the pump 202 speed reduced. The variation intwo-phase quality in line 106 will impact the quality in the flash tank108 and therefore, the liquid level. As the ejector 104 outlet qualityincreases the liquid level in the flash tank 108 will decrease, thusnecessitating closing of valves 208 and 206 and opening valve of 210 toretain a constant liquid level. All of these changes should be made insmall increments, slowly, as the combination of charge, temperature, andpressure propagation throughout the system 300 may take time to settleand can become unstable if large adjustments are made too quickly.

The pressure-enthalpy (P-h) diagrams in FIGS. 4-6 illustrate thepotential performance of the ejector refrigeration cycle system in FIG.2 . The notable feature is the transition between state 3 and state 5 oneach plot. This transition thermodynamically illustrates how a smallincrease in pressure, shown by states 3 and 4, can allow the ejector 104to operate in such a way that a lower stage compressor is not requiredto facilitate stable operation of the refrigeration cycle formulti-stage condensing. The transition between state 3 and 5 alsoillustrates the benefit of employing a pump 202 in combination with theejector 104. By doing so, state 5 is not located close to the saturatedliquid line, as would be expected in a typical vapor compression cycleand is instead brought closer to the saturated vapor line. While FIGS.4-6 are logarithmic pressures—specific P-h diagrams used to illustrateprojected operation of certain working fluids of the ejectorrefrigeration cycle system in FIG. 2 —this does not preclude otherrefrigerants from being used in place of the illustrated refrigerantsfor the system.

While the present disclosure has been described in connection withpresently preferred embodiments, it will be understood by those skilledin the art that it is not intended to limit the disclosure to thoseembodiments. Preexisting ejector refrigeration cycles may be retrofittedor modified according to the disclosure herein, which may also beimplemented in any other refrigeration process employed in an enclosedstructure for heating or cooling to achieve similar results. It istherefore, contemplated that various alternative embodiments andmodifications may be made to the disclosed embodiments without departingfrom the spirit and scope of the disclosure defined by the appendedclaims and equivalents thereof.

1. A refrigeration system, which comprises: an ejector in fluidcommunication with a first evaporative heat exchanger, a pump and aflash drum, the ejector positioned downstream of the first evaporativeheat exchanger and the pump and positioned upstream of the flash drum; asecond evaporative heat exchanger in fluid communication with the flashdrum and a single compressor; the compressor in fluid communication withthe flash drum and positioned downstream of the flash drum; and acondensing heat exchanger in fluid communication with the compressor andthe pump, the condensing heat exchanger positioned downstream of thecompressor and upstream of the pump.
 2. The refrigeration system ofclaim 1, further comprising an expansion valve in fluid communicationwith the condensing heat exchanger and the flash drum, the expansionvalve positioned downstream of the condensing heat exchanger andupstream of the flash drum.
 3. The refrigeration system of claim 1,further comprising at least one sensor for detecting a temperature and apressure in a line connecting the pump and the ejector.
 4. Therefrigeration system of claim 1, further comprising at least one sensorfor detecting a temperature and a pressure in a line connecting thecondensing heat exchanger and the pump.
 5. The refrigeration system ofclaim 1, further comprising at least one sensor for detecting atemperature and a pressure in a line connecting the compressor with theflash drum and the second evaporative heat exchanger.
 6. Therefrigeration system of claim 1, further comprising at least one sensorfor detecting a temperature and a pressure in a line connecting thefirst evaporative heat exchanger and the ejector.
 7. The refrigerationsystem of claim 2, further comprising another expansion valve in fluidcommunication with the flash drum and the compressor, the anotherexpansion valve downstream of the flash drum and upstream of thecompressor.